Camshaft for internal combustion engine valve gear



y 2, 1957 M. c. TURKISH 3,316,890

CAMSHAFT FOR INTERNAL COMBUSTION ENGINE VALVE GEAR Filed p 2, 1956 13 Sheets-Sheet 1 LOADED AND DEFLECTED CAMSHAFT INTAKE CAM EXHAUST cm 4 3 2 I Fl G 2 P2=VALVE SPRING LOAD ON INTAKE CAM s DEFLECTION 0F SHAFT AT EXHAUST CAM FIG. I QIIQ UNDESIRED com/ irm. THE PROBLEM UFTER L AM 7 M OPENED EXHAUST CLOSED H65 3 FIG.4

/INTAKE VALVE EXHAUST CAM LIFTER COMPENSATED BECAUSE I OF EARLIER DEFLECTION AFTER CAMSHAFT DEFLECTION IS L REMOVED.

qNTAKE VALVE SEATED EXHAUST VALVE SLIGHTLY OPENED BECAU o u E [NV TOR F FT R SEIQEEQITIYONAFIEEEZEMOVED EE SE ALVE SHOULD MICHAEL C'TURKISH FIG. 7 FIG. 8 BY J ATTORNEYS May 2, 1967 M. c. TURKISH 3,316,890

CAMSHAFT FOR INTERNAL COMBUSTION ENGINE VALVE GEAR Filed p 1966 15 Sheets-Sheet 2 THE INVENTION LIF ERS IO 4 2 FIG. 9

|NT2NIE E Q EXHAUST TAKE CYLINDER 4 IO FIG. I2 64 EXHAUST'IB so INTAKE CYL.3 GYM 9 FIG. I0 I 62 EXHAUST e2 EXlgAUST 3 75 THE INVENTION\I /-\CONVENTIONAL CAM PROF'LES DEPRESSED PORTION 0N CAM HEEL 6s |NTAKE-CY| 4 WITH BLEND TO LOBE PROFILE ON EXHAUST CAMS 62 AND 64.

CI-INTAKE \z CYL. 3

REPRODUCTION OF ACTUAL OSCILLOGRAM SHOWING PRESENCE OF EXHAUST VALVE ANGLE FORCING CAM BASE CIRCLE REPRODUCTION OF ACTUAL OSCILLO- GRAM SHOWING ABSENCE OF EXHAUST He II VALVE FORCING BY CAM HEEL.

FI I3 TYPICAL CAM BASE G CIRCLE FORCING COMPOSITE OSCILLOGRAM PROBLEM. ILLUSTRATING IMPROVEMENT E TO THE INVENTION. I

INVENTOR.

MICHAEL C. TURKIS H BY ATTORNEYS IM PROVEMENT BY THE INVENTION.

15 Sheets-Sheet 4 FIG. 2|

INVENTOR.

MICHAEL C. TURKISH ATTORNEYS ITHIN BEARING SPAN) M. C. TURKISH CAMSHAF'T FOR INTERNAL COMBUSTION ENGINE VALVE GEAR Filed Sept. 12, 1966 EXHAUST CAM PROFILED HEEL (FOR 2 CAMS W P VALVE SPRING LOAD ON INTAKE CAM S =DEFLECT|ON 0F SHAFT AT EXHAUST CAM May 2, 1967 DEFLECTED CAMSHAFT INTAKE CAM LOAD 8 EXHAUST CAM DEFLECTION 2 2 O m F m N I O 0 I 8 m m mm n O QZEmW m T m 0 S R v 0 w 0 0 ED Us I I 6 V AN D O 3 RmwIdEI IwI XP 8 9 UE O N E m N m fins Fr mm M m n m LoIqoa vufiwo F A E R c G wN m L E E C D G I Um A C D E T F T T IM A VAU .AKU O m M H Err 9 I X H T A E G I S C N W/ P U I M E M I o J. T hm: E U nq s NU A 0 TH ZTZOFOU EMQ O MW 0 V/ 0 m N/ P M IAN W C E S E LR S K FU E A EU W M T A m M D W Ca CAM DEGREES FIG. 24

M. C. TURKISH May 2, 1967 3,316,890 CAMSHAFT FOR INTERNAL COMBUSTION ENGINE VALVE GEAR 13 Sheets-Sheet 6 Filed Sept. 12, 1966 on @E E 01 m. m MN.

ATTORNEYS M. C. TURKISH May 2, 1967 l3 Sheets-Sheet 7 Filed Sept. 12, 1966 mwmmwwm 354 2% NM QE H ow oi owm om 0mm 8m 3w SN 03 om. m. 03 ON. 09 om ow ow 8 0mm I M n X n T QNJQEEE 0mm wofimsw 1 wkzwwumnmm fizmmwmmmm omQL E A H m 8. 952 No- (25m oom w ON.H

0mm 00 52:5 Now. T/ OZ aoo. now. 5 @m OE mm 6E ATTORNEYS M. C. TURKISH May 2, 1967 CAMSHAFT FOR INTERNAL COMBUSTION ENGINE VALVE GEAR Filed Sept. 12, 1966 15 Sheets-Sheet a LOCALIZED WEAR ON CAM HEEL IN REGION WHERE CAM FOLLOWER LOADS HEEL.

CAM L'OBE LOCALIZED WEAR IN CENTRAL REGION OF 253 CAMSHAFT CENTER 8 H M A E H I M m I! 0 T T E Q Q U 2 D R E w u. G mm mm Cw CLOSED INTAKE OR EXHAUST VALVE.

HIGH PRESSURE ON VALVE HEAD AND CYLINDER HEAD DURING COMBUSTION CYCLE CAUSES VALVE ELONGATION AND SHORTENING OF F l G 38 DISTANCE '0 BETWEEN VALVE SEAT AND CAMSHAFT CENTER.

INVENTOR. MICHAEL C. TURKISH ATTORNEYS M. c. TURKISH 3,316,890 CAMSHAFT-FOR INTERNAL COMBUSTION ENGINE VALVE GEAR May 2, 1967 15 Sheets-Sheet 9 Filed Sept. 12, 1966 9 3 G mm c O Z u FMDQIXM II F M O O P 0 R 3 5 O T. O m A w I M S a X G C E E W 6 n a K m a O D R A f D H o E O T L A m m L M 1/ B L A W C M H A 0 S S J w m m w M o A B C F A O R G E 3 L U M w 0 O o O 0 0 w 5 o 0 CRANK ANGLE-DEGREES. T.D.C.

A 9 3 M $3 96 23 5356 m 0 0 0 0 w F O 3 IT 08 s w M w L m &-S .1 L m E m m X non J! D 3 0 w u" E R- J m H S H m m M E 0 H! J B M m III Flm T S A M A PM W n w L A m h I. v E C l v A w F V W o mu M M mm G V 3N o E Mv. A L K J K 1. m L N w w o w w o a s 3 mQ'TQQOJ 2 0 mxqkvz mm mmzmmmmn mmmz3 u ATTORNEYS y 2, 1957 M. c. TURKISH 3,316,890

CAMSHAFT FOR INTERNAL COMBUSTION ENGINE VALVE GEAR Filed Sept. 12, 1966 EYLINDER SEAT DEFLECTION =C FIG. 40 VALVE DEFLE CTION v X= DISTANCE TIP TRAVELS TOWARD J LIFTER'BI CAM HEEL T DURING COMBUSTION.

SEAT EFLECTION X=C+v 13 Sheets-Sheet 10 -TOTAL DEFLECTION INTAKE VALVE FIG. 4|

DE FLECTION TOTAL DEFLECTION EXHAUST NOTE DEFLECTION I. DEFLECTI ON E.

SINCE INTAKE VALVE HAS MORE FLEXIBLE HEAD FIG. 42

DIAGRAM OF CYLINDER PRESSURE AS PHASE RELATED TO CAM HEELS CAM HEEL PROFILING REQUIRED FOR INTAKE 8| EXHAUST DUE TO SHOR TENED CAM Q TO VALVE SEAT DISTANCE RIGID CAMSHAFT INTAKE CAM LOBE ANGLE BETWEEN CAM LOBES INVENTOR. EXHAUST CAM LOBE MICHAEL umqsu FIG. 43 a ATTORNEYS May 2, 1967 Filed Sept. 12, 1966 M. c. TURKISH 3,316,890

CAMSHAFT FOR INTERNAL COMBUSTION ENGINE VALVE GEAR l3 Sheets-Sheet 11 SOLUTION SUITAB LY PROFILE THE CAM HEELS AS SHOWN ON THE FOLLOWING GRAPH. CAM HEELS TO HAVE DEPRESSED REGION WHERE LIFTER CONTACTS HEEL DURING COMBUSTION INTERVAL.

I CYLINDER coMBUsT'IoN PRI-:ssURI-: AT FUIJ.

A THFOTT E I 0 I80 360 540 360 I80 T DIIC. CRANK ANGLE 540 I I I T MAXIMUM INTAKE CAM IN AKE CAM /I IFT HEEL PROFI'UNG' l I I EXHAUST 0AM I g LIFT m I l i I I I L MAXIMUM I EXHAUST I CIAM HIEEL FROFIILE I 270 I80 90 o 90 I I80 270 CAM ANGLE NOTE= CAM HEEL PROFILE sI-IAPEa AMOUNT RELATED TO CYLINDER PRESSURE AT MAXIMUM COMBUSTION PRESSURE;

CYLINDER HEAD BI SEAT DEF'LECTIONS; INTAKE SIEXHAUST VALVE HEAD DEFLECTIONS',

IN SUM MARY= TOTAL VARIATION (SHORTENING OR LENGTHENING) OF THE DISTANCE FROM SEATIIN CYLINDER HEAD TO CAMSHAFT CENTERLINE.

FIG. 44

I I NVENTQR. MICHAEL CI TU RKISH BY ATTORNEYS M. C. TURKISH CAMSHAFT FOR INTERNAL COMBUSTION ENGINE VALVE GEAR Filed Sept. 12, 1966 13 Sheets-Sheet 12 VALVE LASH ADJUSTING DEVICE MECHANICAL OR HYDRAULIC m m V m UR IS DIRECT ACTING TAPPET BY C T K H ON OVERHEAD CAMSHAFT REMOVABLE LASH CAP. LASH ADJUSTED BY usms DIFFERENT THICKNESS CAPS.

m LINE VALVES PIC-7.46

ATTORNEYS.

M. C. TURKISH May 2, 1967 CAMSHAFT FOR INTERNAL COMBUSTION ENGINE VALVE GEAR Filed Sept. 12, 1966 13 Sheets-Sheet 15 EXHAUS BIO INVENTOR.

MICHAE L C. BY

TURKISH ATTOR NE Ys United States Patent CAMSHAFT FOR INTERNAL COMBUSTION ENGINE VALVE GEAR Michael C. Turkish, Lyndhurst, Ohio, assignor to Eaton Yale & Towne Inc., Cleveland, Ohio, a corporation of Ohio Filed Sept. 12, 1966, Ser. No. 578,528 20 Claims. (Cl. 123-90) This invention relates to valve gear systems for internal combustion engines; and, more particularly, to camshafts for internal combustion engines.

Still more particularly, this invention relates to cam profiles that minimize valve forcing phenomena, thereby overcoming valve burning and valve seat sinking resulting from lifter compensations produced by camshaft deflections due to valve spring loadings on adjacent cam lobes.

Also more particularly, this invention relates to cam profiles that eliminate abnormal wear on valve lifters, including rocker arms used as such, produced by deflections of yieldable portions of valve gear in rigid camshaft engines.

This application is a continuation-in-part of copending application Ser. No. 452,669, filed May 3, 1965, and now Patent 3,272,189 dated Sept. 13, 1966.

The problem In any internal combustion engine, precise valve openings and closings are necessary in order to provide proper engine performance and maintain long valve life.

The valves must open and close at exact intervals and must not crack open prematurely or remain cracked open after the proper closing time. If this happens to an exhaust valve, it will quickly burn and its valve seat will sink as the hot burning gases sweep past. Also, time for heat dissipation from the valve head to the valve seat is decreased. If this happens to an intake valve, it will cause idle roughness, impaired fuel economy and diminished valve life because of the disturbance of the precise intake timing event.

Individual cams are provided in an internal combustion engine to impart the properly timed motion to the various intake and exhaust valves. These cams are made part of an integral camshaft which is provided with suitable journals for rotation in closely fitted bearings. The camshaft is geared to the crankshaft to provide synchronous and properly phased rotation.

Since engine performance is dependent upon valve lift, proper event timing, event length and shape of the lift curve, these factors are generally compromised to give the most desirable high speed performance and still retain reasonably good engine idling qualities. A cam lobe profile is designed to provide the proper valve lift motion with due consideration being given to the various parameters of lift, velocity and acceleration. Opening and closing ramps are also provided at the beginning and end of the cam lobe profile to open the valve without impact, to seat the valve slowly to prevent bounce, and to eliminate valve seating noise.

A most important component of the cam is the base circle. This has been treated by the prior art as a nonfunctional component which merely permits the valve to remain closed for an appropriate interval. Previously, there has been no function in this portion of the cam to account for camshaft deflections. Thus, in the prior art, this portion of the cam has been referred to as a base circle portion because it had a uniform radius throughout its length.

By employing a base circle, it is theoretically possible to add a self-compensating lifter to remove the clearance that normally appears in a valve gear under the conditions of a theoretically rigid camshaft and valve gear components, including the valve head and the valve seat portion of the engine head. However, the self-compensating feature causes trouble when camshaft deflections enter the picture.

Camshaft deflections can produce interactions between adjacent valves and cause improper opening. As mentioned above, valve burning and valve seat sinking are the result on exhaust valves. Idle roughness and other malfunctions are the result on intake valves.

In many instances, it is not economically desirable to produce an absolutely rigid camshaft that is free of deflections at all engine speeds. To do so requires a journal adjacent to each cam. Even with an expensive, rigid camshaft construction, valve gear bounce at high speed operations can still be encountered.

Accordingly, in production camshafts, it is one common practice to position a plurality of cams between a pair of spaced journals. A specific example is a typical production camshaft for an in-line six-cylinder engine having four cams between journals. Thus, two central cams are located side-by-side at mid-span and the two end cams are located adjacent to the journals. In this situation, the cams often produce mechanical interactions with one another as the result of camshaft deflections produced by the valve spring loadings on adjacent cam lobes.

Particularly in engines utilizing hydraulic valve lifters, improper valve closing has been encountered at both low and moderate engine speeds. The hydraulic lifter is employed in a valve gear system to remove auto matically any existing clearance appearing during the base circle interval. This, theoretically, serves to produce a lash-free system so that when the cam lobe rotates to the position of causing the valve to lift again, the valve immediately follows the cam lobe without any undue lag or impact. However, the existence of mechanical interaction between adjacent cams, due to camshaft deflections, produces the phenomenon of hydraulic lifter fill-up which causes undesirable compensation of the lifters. This results in the valves being forced just off of their seats during certain portions of their normal valve closed intervals.

Undesirable hydraulic lifter compensation occurs during a part of the base circle interval, in instances where camshaft deflections result from the loading of the adjacent cam lobes. Depending upon the angular relationship between the loaded cam lobes and the cam which has thehydraulic lifter on its base circle, some forcing of a particular valve from its seat may be experienced.

This problem is illustrated in FIGURES 1-8 wherein consideration is given to low and medium speeds of operation. Undesirable conditions occur when high load valve springs are used and when the camshaft has a measurable amount of flexure due to the high loadings caused by the springs.

Thus, as shown in FIGURES 1-3, a valve spring load P imposed on the peak of one cam lobe, intake cam I for example, in its valve-open position, produces a maximum deflection of that cam. But, this is not objectionable since that valve is open and the lifter is solidly loaded. The deflection is shown greatly exaggerated between the dotted and solid lines. However, this deflection is transmitted by the camshaft to an adjacent cam, exhaust cam E for example, which is in a valve-closed position. This causes the valve-closed position cam E to drop slightly, tending to move the base circle away from the face of the hydraulic lifter.

S is the deflection of the exhaust cam. Since the exhaust cam E is closed, the deflection S imposes no harm on the exhaust valve, merely permitting the spring of the exhaust valve to hold it in a closed position. However, the lifter senses the clearance so produced and fills up to remove the clearance S Thus, contact is maintained between the face of the lifter and the base circle of the cam E.

After a hydraulic lifter has compensated for the clearance S introduced between the lifter face and cam E, FIGURE 4, due to the deflection caused by the loading of the cam lobe I, it will remain filled; that is, undesirably compensated upon further rotation of the camshaft when the pre-existing peak spring load on the first cam lobe I is relieved and the camshaft reverts from the deflected condition to the straight, undeflected condition. Since the hydraulic lifter on the adjacent cam E has compensated, it will force its valve slightly off its seat during a part of the base circle interval at the instant when the camshaft reverts to the undeflected condition.

This is the precise point at which the trouble arises. Thus, in FIGURES 5 and 6, the camshaft has rotated clockwise by approximately 120 to the position where both valves are theoretically closed. As shown in FIG- URE 7, the intake valve is closed and the deflection has been removed as indicated by the absence of any dotted line.

However, as shown in FIGURE 8, the prior undesired lifter compensation S is still present for the exhaust valve gear. This causes the exhaust valve to be slightly lifted off its seat as the lifter rides the base circle prior to the opening side of the cam lobe. At this position, the exhaust valve should be fully closed instead of encountering the slightly open condition. This condition causes improper valve functioning.

This undesirable lifter compensation may occur during any part of the base circle interval and will depend upon the amount of greatest camshaft deflection.

If the greatest amount of camshaft deflection occurs in the beginning part of the valve-closed base circle interval, undesired lifter compensation will occur there; that is, just following the normal lift cycle. Then, however, in the latter part of the base circle interval, the deflection disappears; but, the lifter remains undesirably compensated. This causes the valve to be forced open during the latter part of the base circle interval, or just prior to its normal lift cycle.

Similarly, if the greatest amount of deflection occurs in the latter part of the valve-closed base circle interval, undesired lifter compensation will occur there; just prior to the normal lift cycle. Then, however, after the valve lift cycle is completed, the deflection disappears; but, the lifter remains undesirably compensated. This causes the valve to be forced open during the early portion of the valve-closed base circle interval, just after the valve lift cycle.

The very limited amount of lifter leak-down that occurs during each cycle, when the lifter is loaded on the lift portion of the cam lobe, is only one thousandth of an inch (.001) or less. This is insuflicient to offset the effect of lifter compensation occuring when deflections exceed this amount. High rates of lifter leakage, by use of large clearances or low oil viscosities, will allow a greater tolerance to camshaft circle deflections. But, this is not a satisfactory solution since it results in noisy and erratic valve seating.

Self-adjusting mechanical lifters operating with zero clearance are similarly sensitive to this problem of camshaft deflection. The same problems will also occur with these mechanical self-compensating devices as are encountered with hydraulic lifters Where camshaft deflections cause valves to be improperly forced off of their seats.

Mechanical tappets, which have no self-adjusting means and have a reasonably large appropriate clearance provided within the valve gear, are not so sensitive to this problem of camshaft deflections. Positive valve closing can be assured when the proper valve gear clearance is provided. However, if the valve gear is designed to operate with very low clearance and the camshaft is capable of large deflections, in excess of the clearance provided, then the forcing of the valves off of their seats, as encountered with hydraulic lifters, can occur.

Thus, at low and moderate engine speeds, valve malfunctions are often produced by mechanical interactions of adjacent cams located between the same pair of journals. Spring loads imposed on the cam lobes are effective to deflect the camshaft and thus produce the cam base circle valve-forcing phenomena, caused by lifter compensations occuring during these deflections.

Therefore, a substantial and important advancement to the art would be provided by an improved cam profile that would counteract undesired hydraulic lifter compensation during the base circle interval to alleviate improper opening of valves of internal combustion engines using hydraulic lifters.

Further, an important advancement to the art would be provided by a novel cam profile that would counteract undesired lifter compensation produced by camshaft deflection to alleviate improper opening of valves of internal combustion engines using self-adjusting mechanical or similar lifters.

Still further, an important advancement to the art would be provided by a novel and improved cam profile operating with valve gear set to low clearances, that would alleviate improper opening of valves of internal combustion engines using mechanical tappets.

Further elements of the problem:

In addition to the aspect of camshaft deflection which has been analyzed in detail above, there are elements of the valve gear, as distinguished from the camshaft, which can deflect at engine operating speeds to produce malfunctions. This presumes that the camshaft can be manufactured in a manner to be rigid so that the problems encountered with camshaft deflections are theoretically not present.

For the present analysis, this then means that other elements of the valve gear may deflect. Thus, a rather large valve head, as used in modern engines to assure free breathing, is in the nature of a thin shell that will bow under combustion pressures and deflect the valve stem axially as in a direction away from the base of the valve head. Also, it is a recognized phenomenon that portions of the cylinder head above the combustion zone, namely the valve seat and surrounding portions, will deflect under combustion pressures.

Thus, deformations of the valve and/ or deflections of the cylinder head (engine which includes the valve seat) impose a loading on the valve gear. These deflections produce an effective shortening of the valve seat to cam distance. Essentially then, the cam base circle becomes loaded during a substantial portion of its interval. This causes the oil film to be scrubbed away and results in localized intensive wear on the follower.

Hydraulic lifters must compensate for deflections of these yieldable portions of the valve gear in order to prevent abnormal wear on the lifters. However, in the case of hydraulic lifters, this distortion phenonrnenon shortens the time for the lifters to compensate. only noise-but, more importantly, localized wear on the follower where the base circle of the cam profile is forced into scrubbing contact with the follower-and scrubs away the oil film. The resultant wear disturbs the whole valve action and requires engine rebuilding to alleviate the damage and restore the engine to proper operating condition. Such restoration, however, is short lived and the process repeats itself. Therefore, the cure is not permanent.

Accordingly, a still further substantial advancement to the art would be provided by an improved cam profile The result is not.

that would prevent undesired alleviate abnormal follower wear.

Objects, therefore, are to provide novel cam profiles that are effective to minimize valve forcing phenomena produced by camshaft deflections.

Further objects are to provide novel cam profiles that are effective to eliminate abnormal wear on valve follower surfaces produced by deflections of yieldable portions of the valve gear in rigid camshaft engines.

FIGURE 1 is a schematic elevational view of a portion of a camshaft having two cams located within a bearing span, illustrating the problem that has been solved in accordance with the present invention;

FIGURE 2 is a schematic sectional view taken along the line 22 of FIGURE 1;

FIGURE 3 is a schematic sectional view taken along the line 3-3 of FIGURE 1;

FIGURE 4 is a schematic sectional view taken along the line 44 of FIGURE 1;

FIGURE 5 is a view similar to FIGURE 1 but showing the camshaft in a partially rotated position beyond the FIGURE 1 position, where both valves should now be closed;

FIGURE 6 is a schematic view taken along the line 6-6 of FIGURE 5;

FIGURE 7 is a schematic View taken along the line 7-7 of FIGURE 5;

FIGURE 8 is a schematic view taken along the line 8-8- of FIGURE 5 FIGURE 9 is a schematic elevational view, partly in section, of a portion of a camshaft wherein four cams are located within a bearing span, to provide an environment for discussion of the invention;

FIGURE 10 is a schematic end-elevational View, taken along the line 10-10 of FIGURE 9, illustrating conventional cam profiles;

FIGURE 11 is a reproduction of an actual oscillogram of a load-time or load-angle diagram, showing the presence of exhaust valve forcing while on the base circle, produced by the conventional cam profiles of FIGURE 10;

FIGURE 12 is a schematic view illustrating one aspect of the present invention;

FIGURE 13 is a reproduction of an actual oscillogram of a load-time or load-angle diagram, showing the absence of exhaust valve forcing while on the novel functional heel resulting from use of the invention as shown in FIGURE 12;

FIGURE 14 is a composite oscillogram, made by superimposing the small circled portions of FIGURES 11 and 13, to illustrate more clearly the improvement obtained by use of the present invention;

FIGURE 15 is an illustration of a segment of a theoretically rigid and undeflected camshaft having two cams within a bearing span, for purposes of further illustrating the problem;

FIGURE 15:: is a schematic view taken along the line 15a-15a of FIGURE 15;

FIGURE 16 includes a graph illustrating normal lift curves for exhaust and intake cams of the undeflected camshaft of FIGURE 15, but with deflection shown by the dotted line;

FIGURE 17 illustrates the camshaft deflection encountered when two cams are located within a bearing span, wherein spring load on the exhaust cam lobe produces deflection of the intake cams;

FIGURE 18 is a graph showing intake cam deflection as related to exhaust cam lobe loading;

FIGURE 19 illustrates the principle of the invention as applied to profiling the heel of the intake cam by depressing a portion of the heel in its early portion;

FIGURE 20 is a graph showing the function of the novel profiled heel of FIGURE 19 to prevent improper forcing of the intake valve off of its seat as a result of camshaft deflection, as illustrated in FIGURE 17;

FIGURE 21 illustrates the camshaft deflection encounloading on the valve gear to tered with two cams within a bearing span, wherein valve spring loading on the intake cam lobe produces deflection of the exhaust cam;

FIGURE 22 is a graph showing exhaust cam deflection as related to intake cam loading;

FIGURE 23 illustrates the principle of the invention as applied to profiling the heel of the exhaust cam by depressing a portion of the heel in its latter portion;

FIGURE 24 is a graph showing the function of the novel heel profile of FIGURE 23 to prevent improper forcing of the exhaust valve as the result of camshaft deflection, as illustrated in FIGURE 21;

FIGURE 25 is an illustration of three cams within a bearing span as used in an opposed-piston engine, for purposes of showing exhaust cam deflection;

FIGURE 26 is a schematic view taken along the line 26-26 of FIGURE 25;

FIGURE 27 is a graph showing deflections of two exhaust cams as the result of valve spring loadings on the intake cam lobes positioned therebetween, as in FIG- URE 25;

FIGURE 28 illustrates the principle of the present invention as applied to profiling the heel of the exhaust cams of FIGURE 25;

FIGURE 29 is a graph showing the function of profiling the heels of both exhaust cams in FIGURE 25, in accordance with FIGURE 28, to prevent improper forcing of the exhaust valves of FIGURE 25 FIGURE 30 is an illustration similar to FIGURE 25, of three cams within a bearing span as used in an opposedpiston engine, for purposes of showing opposed intake cam deflections;

FIGURE 31 is a schematic view taken along the line 31-31 of FIGURE 30;

FIGURE 32 is a graph showing deflections of the intake cam of FIGURE 30, as the result of valve spring loadings on the intake cam lobe and on the two adjacent exhaust cam lobes;

FIGURE 33 illustrates the principle of the invention as applied to profiling the heel of the intake cam of FIG- URE 30;

FIGURE 34 is a graph showing the function of profiling the heel of the intake cam in accordance with FIG- URE 33; W

FIGURE 35 is a sectional view of a camshaft illustrating the profile of an exhaust cam wherein undesired lifter compensation is prevented in accordance with the principles of the invention, by elevating a portion of the novel profiled heel above a theoretical base circle radius to provide a function equivalent to the previous showings wherein a portion of the heel was contoured to a level below a theoretical base circle periphery;

FIGURE 36 is a sectional view similar to FIGURE 35, but showing an adjacent intake cam profile, wherein undesired lifter compensation is prevented in accordance with the principles of the invention, by elevating a portion of the novel profiled heel above a theoretical base circle radius, to provide a function equivalent to the previous showings wherein a portion of the heel was contoured to a level below a theoretical base circle periphery;

FIGURE 37 is a graph illustrating typical exhaust and intake cam lift curves produced by the novel profiled heel cams of FIGURES 35 and 36, for use with a deflected camshaft as encountered in actual engine operation;

FIGURE 38 is a fragmentary sectional view of a production six-cylinder in-line engine having a rigid overhead camshaft, wherein combustion pressures produce valve geardistortions, producing malfunctions;

FIGURE 39 is a reproduction of a typical oscillogram, showing the presence of intake and exhaust valve gear loadings, produced by conventional cam profiles in the engine of FIGURE 38 when operated at low speed and open throttle;

FIGURE 39A is a reproduction of a typical oscillogram similar to FIGURE 39 except that the engine is operated at high speed and open throttle;

FIGURE 40 is a schematic view illustrating the amount of valve seat deflection, of the engine of FIGURE 38, that is taken into consideration in arriving at the profiling, by invention, as shown in FIGURE 43;

FIGURE 41 is a schematic view illustrating the com- 'bined amount of valve seat and intake valve head deflection, of the engine of FIGURE 38, that is taken into consideration in arriving at the profiling, by invention, as shown in FIGURE 43;

FIGURE 42 is a schematic view illustrating the combined amount of valve seat and exhaust valve head deflection, of the engine of FIGURE 38, that is taken into consideration in arriving at the profiling, by invention, as shown in FIGURE 43;

FIGURE 43 is a schematic view showing the cam heel profiling as applied to the engine of FIGURE 38, and illustrating the manner in which the profiling is phase related to cylinder combustion pressure;

FIGURE 44 is a graph illustrating the effect of application of the cam base circle segment profiling, by invention, to the engine of FIGURE 38, to produce improved operation;

FIGURE 45 is an illustration of an engine valve gear, having a rigid overhead camshaft and using either a mechanical or hydraulic lash adjuster, in the nature of the engine of FIGURE 38, to which the principles of the invention are applicable;

FIGURE 46 is an illustration of an engine valve gear having a rigid overhead camshaft and using a direct mechanical tappet, providing position clearance during the valve-closed interval to which the principles of the invention are applicable;

FIGURE 47 is an illustration of an engine valve gear, using a rigid overhead camshaft with mechanical tappets for the intake valves and a rigid block-mounted camshaft with hydraulic lifters and rocker arms for the exhaust cams, to which the principles of the invention are applicable; and,

FIGURE 48 is an illustration of an alternate arrangement of the engine of FIGURE 47, wherein a single, rigid, overhead camshaft is used for both sets of valves, mechanical tappets being used for the intake valves, and hydraulic lifters and rocker arms for the exhaust valves, to which the principles of the invention apply.

The terminology profiling the heel refers to providing a profile for the heel of the cam which is essentially nonuniform radially as distinguished from a base circle of the prior art. In one embodiment of the invention, this is effected by depressing selected regions of the heel to counteract camshaft deflections. In an alternate embodiment, this profile can be described as elevating selected regions of the heel to counteract camshaft deflections. In each case, the precise end result of the profiled heel is the same. Therefore, the terminology profiled heel is adhered to and encompasses both embodiments.

Actually, this invention was carried out by modifying existing cams having base circles of uniform radii, employing the first embodiment to prove the merits of the invention in actual engine tests. This shows that within the scope of the invention, profiled heels can be generated per se, or by undercutting conventional base circles in either the same or separate grinding operations.

Indeed, the invention can be advantageously applied to a cam design by specifying the lift figures to five decimal places, as in the usual practice, for each degree of cam rotation throughout the entire 360. In this manner, both the heel profile and the entire cam lobe profile are specified as one continuous cam lift curve. In this manner all vagueness of terminology is eliminated and the true purpose of this invention can be precisely and systematically applied.

8 The invention; FIGURES 9I4 In accordance with the present invention, instead of a base circle, a novel, profiled heel is employed. Thus, lifter compensation which occurs during camshaft deflection is subsequently counteracted and as a result the value is kept properly closed during the heel portion of cam operation.

In FIGURE 9, there is illustrated a portion of a camshaft 50. Bearings 52 and 54 support the camshaft at spaced intervals along its length, establishing the bearing span. The bearings 52 and 54 receive journals 56 and 58 respectively. Within the bearing span are four cams 60, 62, 64 and 66. Engaging each of the cams 60, 62, 64 and 66 are lifters 68. Each lifter is operably connected to a valve.

The element 70 is a fuel pump eccentric frequently located centrally of a six-cylinder engine camshaft, as illustrated.

The camshaft illustrated is actually used with an overhead valve, six-cylinder engine. In this arrangement, each of the lifters 68 is operably connected to the bottom end of a pushrod that extends up along the side of the engine to engage one end of a rocker arm journaled on a rocker shaft, bolted to the top side of the head of the engine. The other end of each rocker arm engages the end of a valve stem. Each valve stem in turn carries a valve spring, not shown, held captive between the top side of the engine head and a valve spring retaining washer attached to a groove on the valve stem. The valve spring thus is effective to continuously bias the valve toward a closed position, in turn forcing the rocker arm in a manner to force a lifter 68 downwardly to contact its cam, as shown in FIGURE 9.

The lifters 68 are self-adjusting hydraulic lifters and, theoretically, maintain constant contact with the cams so that the-re is no clearance or lash in the valve gear to produce noise, vibration and impact shock loads.

While the foregoing description has related to valve-inhead engines, the same principle applies to valve-in-block engines.

In FIGURE 10, a conventional intake cam 60, for cylinder #3, is shown as having a true radius 72 as the base circle on the valve-closed portion of the cam. Typical of the prior art, this is nonfunctional. The cam also has a lobe 74. Typical of the prior art, this is the only functional portion.

In the condition of FIGURE 10, which is the same as shown in FIGURE 9, the lobe 74, having spring-loaded lifter 68 on its peak, is that of the intake cam 60. In this condition, the intake valve of cylinder #3 is open. This imparts maximum spring load on the camshaft at that point and causes a downward camshaft deflection similar to that illustrated by FIGURE 1. This deflection at cam 60 is also imparted to the adjacent cam 62 which, in this instance, is the exhaust cam for cylinder #3.

Now visualize that the camshaft rotates as indicated by arrow 76. The lifter 68 rides down the lobe 74 in the arrow direction 78 so that the valve seats when the base circle 72 reaches the top side of the camshaft. The deflection caused by intake cam 60 decreases as the spring load is reduced, while the valve of cam 60 approaches its seat; and then, the deflecting load source finally is eliminated as the valve becomes seated. Another source of deflect-ion producing loading occurs as exhaust cam 64 operates its lifter and valve. In similar manner, the valve spring load acting on cam 64 produces a downward deflection to the camshaft even after cam 60 rotates beyond the point where it causes no deflection. Now as lifter 68 rides down the lobe 75 of cam 64, the deflection caused by this cam decreases as the spring load is reduced. During this second unloading of the camshaft by cam 64, the camshaft also returns toward its normal undeflected condition.

Now consider the action of cam 62 in FIGURE 9, which is adjacent to both cams 60 and 64.

If, in the angular sector 80 of the base circle, the cam 62 is deflected downwardly by valve spring loads on earns 60 and 64, as has just been explained, the lifter for cam 62 will undesirably compensate.

Then, as the angular sector 82 of the base circle is brought around to engaging position with the lifter face for cam 62, the previous compensation, being present at the instant when the camshaft deflection is relieved, causes forcing of the exhaust valve, being operated by cam 62, and the valve is cracked off its seat. This results in burning of the exhaust valve and sinking of the valve seat due to the leakage of the hot burning gases past the valve and valve seat.

FIGURE 11 is a reproduction of an actual oscillogram showing the presence of exhaust valve forcing by cam 62 of cylinder #3, as shown in FIGURES 9 and 10. This results from the sudden disappearance of the prior deflection when the lifter became filled up.

The invention applied In accordance with the present invention, a novel profiled heel is employed. FIGURE 12 illustrates the manner in which the angular sector 82 of the heel is profiled by depressing its surface to the maximum depth 83 to nullify the forcing condition just described.

FIGURE 13 is a reproduction of an actual oscillogram taken during engine operation, using a novel profiled cam heel as illustrated in FIGURE 12.

FIGURE 14 more vividly illustrates how the base circle forcing problem has been solved in accordance with the present invention. The upper curve represents the forcing from FIGURE 11, the prior art, imposed upon a similar portion of the curve of FIGURE 13, representing the invention. The improvement of the invention is represented by a horizontal line, indicating the absence of exhaust valve forcing.

The foregoing description provides an illustration of the principle of the present invention and illustrates the manner in which the problems caused by camshaft deflections are overcome. Thus, mechanical interaction of adjacent valves, produced by camshaft deflections caused by loads imposed by operating adjacent cam lobes, are counteracted by the invention.

The intake cam profiled heel; FIGURES 15-20 two cams within bearing span In FIGURE 15, there is shown a theoretically rigid camshaft. This is designated 84, and is illustrated as having two cams within the bearing span as distinguished from FIGURE 9, which had four cams. Here, the exhaust cam is designated 86 and the intake cam 88. The spaced bearings are designated 90 and the journals 92.

FIGURE 16 includes a graph of the lift curves for an 'undeflected camshaft, as illustrated in FIGURE 15. As regards the exhaust cam, there is a positive rise from zero lift in approximately 70 of rotation, passing through a maximum slightly before 180 and then back to exactly Zero at slightly before 270. Similarly, there is a uniformly smooth curve for the intake cam, showin positive rise from exactly zero lift at slightly after 180, to a maximum at 270, returning to exactly zero at approximately 360. It will be observed that at no time during the closed period is there any indication of anything other than exactly zero lift. Thus, the valves theoretically are tightly closed during the entire base circle interval.

It can be readily understood that if there was any forcing of the intake cam due to exhaust cam lobe loading and deflection, there would be a partial intake valve opening as indicated by the dotted line 94 in FIGURE 16. This condition would result from the type of deflection shown in FIGURE 17.

Referring more particularly to FIGURE 17, there is a maximum valve spring load P imposed on the exhaust cam 86. This produces a maximum deflection at this point. Also, it produces a deflection S of the camshaft as far away as intake cam 88. This is shown graphically in FIGURE 18. Thus, as shown in FIGURE 18, as the exhaust cam load rises to a maximum, the intake cam deflection similarly reaches a maximum returning to zero when the exhaust cam loading has passed.

In this instance, this deflection has no effect on the exhaust valve because it is in an open position. However, there is a damaging effect on the intake valve. Thus, by reference to FIGURE 17, it is to be understood that a hydraulic lifter 68 has sensed the deflection S of intake cam 88. Under the conditions of FIGURE 17, therefore, the lifter fills up to IE-QSIabllSh contact with the cam 88.

Now refer to FIGURE 19. The intake cam is actually deflected down to the dotted line position 96. The camshaft rotates in the arrow direction 76 to bring the intake cam up to an opening position. The previous, undesired compensation, which the lifter has made, is not damaging at this instant as the valve opens, since it just tends to open the valve a little further than normal. However, when the intake ca m rotates further to the position where the angular sector 98 of the novel profiled heel is uppermost, the depressed portion, as shown in FIGURE 19, is effective to counteract the compensation previously made by the lifter and thus permits the valve to fully close.

As the angular sector 100, having the elevated portion of the novel profiled heel again comes around to the top side, camshaft deflection will again come into play, as by the adjacent cam 86 opening the exhaust valve, to keep the intake valve properly seated, even with the previous compensation of the lifter still existing.

The function of the depressed portion of the profiled heel is illustrated graphically in FIGURE 20. The intake valve closing period occurs from 90 to near 0. The depressed portion in the shaded area allows the intake valve to become properly seated even though the camshaft returns to its undeflected condition during this interval. Since the camshaft remains undeflectecl from 0 and 90 of rotation, it is necessary to depress the heel for this full interval to counteract the lifter compensation which occurred previously when the camshaft was deflected.

On further rotation, cam 36 actuates the exhaust valve, opening at between 90 and imposing a deflection to both the exhaust and intake cams. The exhaust openmg lift and the resulting deflection are shown as dotted deflection S shown in FIGURE 17, causes the undesired lifter compensation to occur so that, on the next passing of the 0 to 90 angular sector 98, the depressed portion must again counteract lifter compensation.

S nmmaly To make sure that the reader clearly understands the problem and the solution offered by the invention, it is simply reiterated as follows. A camshaft deflection causes a hydraulic lifter, Which is riding the cam heel, to fill up. Then, during the interval on the heel when the camshaft deflection is relieved, the hydraulic :lifter will exert a force upon a valve to unseat it. Thus, in accordance with the invention, the heel is profiled by being depressed in the angular sector that is top side when the camshaft is undeflected so as to counteract the compensated condition of the lifter.

The exhaust cam profiled heel; FIGURES 21-24 two cams within bearing span This, of course, produces a maximum deflection of the intake cam with some of that deflection being imparted to the exhaust cam as S By referring to FIGURE 22, the intake cam loading and exhaust cam deflection are correlated. When the intake cam load is at a maximum, the deflection is at a maximum. When there is no intake cam load, the exhaust cam deflection is zero.

In FIGURE 21, the intake cam is shown at maximum valve-open position, causing maximum valve spring force to be applied to the intake cam. This imparts deflection S to the exhaust cam, as shown in FIGURE 23. This deflection is sensed by the hydraulic lifter for the exhaust cam when the exhaust cam has the angular sector 102 of the novel profiled heel of FIGURE 23 uppermost. As the angular sector 104 of the heel comes up, during camshaft rotation, the valve would normally be lifted off its seat as the deflection S was relieved by the intake cam 106 passing beyond the range of the cam lobe contacting its lifter. In order to counteract lifter compensation occurring during deflection, a depressed portion is provided on the heel in the region nearest the opening side of the exhaust cam lobe. This depressed portion is located principally in the angular sector 104 as illustrated in FIGURE 23.

By referring to FIGURE 24, it will be noted that the exhaust cam closes between l80 and 90. At 90, the intake is fully open imposing maximum spring load and, thus, deflection S on the exhaust cam. This causes the exhaust cam to move away from the face of its lifter and the lifter compensates. As the intake closes near the face of the previously compensated exhaust lifter then rides the depressed portion of the heel to nullify the absence of deflection and, thus, keep the exhaust valve from being forced and held at a partially open position. Thus, the exhaust valve is kept at Zero lift until shortly after 90 where the valve opening properly starts again.

Summary The prior dscription has been directed to two cams within a bearing span in order to clearly illustrate the problem and highlight the application of the principles of the invention thereto.

Additionally, the invention is applicable to a camshaft wherein four cams are included within a bearing span as was originally mentioned relative to FIGURE 9, representing the environment of a specific production sixcylinder engine. Problems of camshaft deflection become more complicated with a greater number of cams between bearing spans. However, the principles of application fully apply.

The exhaust cam profiled heel for opposed-piston engines: three cams within bearing span; FIGURES 25-29 The principles of the invention are not only applicable to in-line piston engines, but they are also applicable to opposed-piston engines, such as used in aircraft and in some present-day rear engine automobiles.

In opposed-piston engine structures of this type, it is frequently possible to permit one cam lobe to actuate two lifters located diametrically opposite each other. This eliminates the need of one cam lobe and produces a more compact engine arrangement. A portion of such a camshaft for operation with two opposed engine cylinders, is shown in FIGURE 25. Three cams are provided between the bearing span, and in this arrangement, it is to be noted that cams 122 and 124 each actuate only one lifter. However, the middle cam operates opposed lifters 125 and 127.

It is to be understood that the arrangement is illustrative and opposed exhaust lifters, operated by a single exhaust cam in the mid span flanked by two intake cams, is also within the scope of the invention.

In this arrangement, referring to FIGURE 25, it will 12 be noted that the two exhaust cams 122 and 124 will be deflected by alternate and opposite spring loads imposed against intake cam 126.

The cam profiled heel, in accordance with the invention, to counteract these deflections, is illustrated in FIG- URE 28. The depressed portion of the profiled heel starts after angular sector 128 and grows deeper in sector 130.

The manner in which the depressed portion functions is illustrated graphically in FIGURE 29. As a prelude to the explanation following, it is to be kept in mind that the deflection diagramed in FIGURE 25 is first in one direction and then in a diametrically opposite direction.

During the angular interval between and 180, shown in FIGURE 29, the exhaust valve 122 of cylinder #1 gradually closes. At the end of this closing stroke, the spring load acting upon lifter of intake cam 126 of cylinder #1 comes into play, imposing a camshaft deflection and thereby causing a compensation by the hydraulic lifter 121 of exhaust cam 122. At about 270, where valve 126 closes, the camshaft deflection is relieved. This would normally tend to force exhaust valve 122 prematurely open. However, the depressed portion propitiously provided to the profiled heel of cam 122 in angular sector comes into effect, as shown in FIG- URE '28, thus keeping the exhaust valve 122 closed.

An unusual region exists between 270 and 360, conforming to the latter part of sector 130. Here the deflection of exhaust cam 122 is reversed by the action of an opposite spring force acting on cam 126 caused by opening the intake valve of cylinder #2. This, in effect, doubles the deflection at exhaust cam 122, and is the reason why the depressed portion becomes deeper near the latter part of angular sector 130 in FIGURE 28.

The opposite direction deflection imparted to the camshaft must be counteracted, else the valve will be forced open for this interval on the heel. It will be noted from FIGURE 29 that all deflections are counteracted by the depressed portion of FIGURE 28; therefore, the valve 122 remains properly closed until 360 where the opening properly begins again. In like manner, exhaust cam 124 requires a similar depressed portion on the heel as illustrated in FIGURES 28 and 29.

The intake ca m profiled heel for opposed-piston engines: three cams within bearing span; FIGURES 30-34 The same camshaft section containing three cams within a bearing span, as in FIGURE 25, has been reproduced in FIGURE 30. Again in this arrangement, it is to be noted that cam lobes 122 and 124- each actuate only one lifter, whereas the intake cam lobe 126 operates opposed lifters 125 and 127. A complex deflection of cam 126 is the result.

As shown in FIGURE 32 reading from left to right, opening of intake valve of cylinder #2 by cam 126 acting on lifter 127, imposes an upward camshaft deflection that is partially balanced by the opening of exhaust valve of cylinder #1 by cam lobe 122.

Later, the opening of the opposite intake valve of cylinder #1 by cam lobe 126 acting on lifter 12S imposes a downward deflection adding to the downward deflection caused by cam lobe 122. anced by the subsequent opening of the exhaust valve of cylinder #2 by cam lobe 124. Finally, the reopening of intake valve of cylinder #2 by cam lobe 126 adds to the upward deflection imposed by cam lobe 124.

The net result of the foregoing rather complex loading of the different cam lobes produces a resultant deflection curve at the intake cam 12d, as indicated graphically in FIGURE 32.

The manner in which the heel of the intake cam is profiled in both sectors 132 and 134 is shown in FIG- URE 33. The function of the profiled heel of FIGURE 33 is illustrated graphically in FIGURE 34. V

The intake valve of cyiinder #2 in FIGURE 30 closes This is somewhat "Dalshallow level as the sector 134 progresses.

. quadrant 204, is substantially at shaft deflections, here lifter 13 at 180. The profile of earn 126 for sector 132 illustrated in FIGURE 33, counteracts the effect of the initial deflection imposed from the opposite direction between 180 and 270 by the loading on cam lobe 122 and the later deflection caused by loading cam lobe 126 to open intake valve of cylinder #1. This keeps the intake valve of cylinder #2 at zero lift. The further deflection encountered by the overlap of exhaust cam lobe 122 with the intake cam lobe 126, both operating valves of cylinder #1, begins at about 270 as illustrated in FIGURES 32 and 34. This compound deflection is counteracted by the depressed portion of the profile in about the central part of the heel as shown in FIGURE 33. As the exhaust valve of cylinder #2 then opens by cam lobe 124, the deflection is counteracted somewhat in the opposite direction so that the depressed portion returns to a more Therefore, between 270 and slightly before 360", the changing profile functions to keep the valve at zero lift.

Elevated profiled heel: FIGURES 3537 In these figures of the drawings, a logical extension of the invention to include relatively elevated profiled heel portions is shown. Thus, the broad principle is extended by providing a specific portion of a novel profiled cam heel of the invention relatively higher, in elevation, then another portion of relatively lower elevation. As developed hereinbefore, the relatively depressed portion of the novel profiled heel functions to counteract undesired, but existing, lifter compensation when the lifter is riding on that portion of the novel profiled heel and camshaft deflection has been simultaneously removed. In this logical extension of the invention, undesired lifter compensation is prevented by the expedient of a novel profiled cam heel portion that is relatively elevated to an extent equalling camshaft deflection. Thus, when camshaft deflection takes place, this elevated heel portion in effect retains the lifter in a condition equivalent to that provided by a theoretically rigid camshaft.

Thus, the undesired lifter compensation is prevented. To develop this logical extension further, reference is i made to FIGURES 35, 36 and 37.

FIGURE 35 represents an exhaust cam 200 in front of an intake cam 202, for illustrative purposes, it being assumed that there are these two cam lobes only within a bearing span. Although the novel profiled heel shown will differ somewhat from the heel profile for three or four lobes within a bearing span, and the manner in which the intake and exhaust cams are oriented in such a configuration, FIGURE 35 nevertheless depicts a typical mode of operation.

Relative to the exhaust cam 200, as shown in FIGURE 35, it will be noted that the quadrant 204 in the early portion of the profiled heel is elevated beyond the circle 206 representing a base circle as used in cams of the prior art. In the quadrant 208, which lies in the latter portion of the novel profiled heel of the exhaust cam 200, the operating surface 210, though relatively depressed compared to the relatively elevated surface 212 of the earlier the circle dimension 206.

It will be noted that this is the converse of the embodiments shown hereinbefore wherein a functional portion was developed by depressing a portion below the periphery of the circle. Although the same function is produced in counteracting valve interactions from camcompensation is prevented rather than being nullified after it has taken place as in the foregoing embodiments.

To further develop this description, identifying numerals are applied to FIGURE 36. Thus, the intake cam 202 has a quadrant 214 in the early portion of the heel where the working surface 216 is at the circle line 206. In the latter portion quadrant 218, the working surface 220 is relatively elevated above the circle 206.

Note that the arrangement is the opposite of FIG- URE 35.

In view of the foregoing, the function of the novel profiled heel for the exhaust cam 200 as well as the novel profiled heel of the intake cam 202 can now be developed by reference to FIGURE 37. Proceeding from left to right in FIGURE 37, it will be noted that the exhaust 200 starts to open between and By reference to FIGURE 35, the correlatives are shown. Between 120 and 180 the lobe 201 causes the valve to open to a maximum extent. Between 180 and about 240 there is a maximum closing with the closing ramp being encountered a little beyond 240. The closing terminates just short of 270. Thereupon, the lifter will ride the functional surface 212, which will be uppermost in FIG- URE 35 at this time, for the remainder of the revolution.

The intake action Retracting now to 120, it will be noted that as the exhaust cam 200 proceeds to a maximum open position at 180, camshaft deflection represented by spring loading reaches its maximum, beginning at 120 and becoming maximum at about and remaining so to about 200 where it is rapidly diminished because the lobe 201 passes and the novel profiled heel takes over.

Under maximum spring loading and thus during camshaft deflection, the quadrant 218 of the intake cam 202, FIGURE 36, is uppermost, i.e. in contact: with the lifter. Due to the elevated functional surface 220, camshaft deflection is counteracted and the lifter is not permitted to undesirably compensate, as in the prior embodiments of the invention. Then, in quadrant 214, after the lobe 201 of exhaust cam 200 has passed and maximum spring load and consequent deflection are removed, the lifter can ride the surface 216, FIGURE 36, which is at the circle 206 line, without causing the valve to be improperly forced open, because the lifter is normal.

From the foregoing, it will be understood that the camshaft deflection increment is substantially equivalent to the elevation produced by working surface 220, FIG- URE 36, as represented by increment elevation a in the FIGURE 37.

Continuing to read from left to right in FIGURE 37, it will be understood that the increment elevation b which is shown somewhat greater than a for purposes of illustration, will function in the same manner when the intake lobe 203 is in contact with the lifter, thus causing deflection of the camshaft when the functional surface 212 of quadrant 204 of exhaust earn 200 is in contact with its lifter.

The foregoing, therefore, presents a logical extension of the invention where undesired lifter compensations are prevented by elevating a portion of the novel profiled heel to an extent equal to camshaft deflections produced by valve interactions.

It is to be understood that this extension is applicable to the various engine ramifications discussed herein, including in-line, V, opposed-piston, etc.

The invention is more particularly applicable to undesired hydraulic lifter compensation arising from camshaft deflections. However, .the invention is also applicable to self-adjusting mechanical lifters, operating with zero clearance, which compensate as a result of camshaft deflections.

The invention is also applicable, to a degree, to malfunctions produced by camshaft deflections when using mechanical tappets intended to operate with a small amount of fixed lash. Thus, if mechanical tappets are improperly adjusted with insuflicient lash and if camshaft deflections are excessive, then mechanical interaction between adjacent cams may produce valve loading and opening during -the normally closed interval.

The rigid camshaft engine As a logical extension, the present invention is also applicable to engines in which a very rigid camshaft is 

1. IN AN INTERNAL COMBUSTION ENGINE HAVING A VALVE SEAT AND VALVE GEAR COMPONENTS WHICH DEFLECT UNDER CYLINDER COMBUSTION PRESSURE, THE ENGINE HAVING A SUBSTANTIALLY RIGID CAMSHAFT INCLUDING A CAM HAVING A LOBE AND A HEEL, THE HEEL PROVIDING VALVE CLOSING, THE VALVE GEAR DEFLECTIONS PRODUCING A CHANGED CAM TO VALVE SEAT DISTANCE, AND INCREASED LENGTH OF VALVE GEAR COMPONENTS CONTAINED THEREIN, DURING THE COMBUSTION INTERVAL TO PRODUCE VALVE GEAR MALFUNCTION BY THE CAM HEEL DURING A PORTION OF ITS INTERVAL, THE IMPROVEMENT WHEREIN A PORTION OF SAID CAM HEEL IS PROFILED TO A MAGNITUDE SUBSTANTIALLY EQUIVALENT TO THE DEFLECTION OF SAID VALVE GEAR DURING CAMSHAFT ROTATION, 